Evaporator refrigerant saturation demand defrost

ABSTRACT

A method is disclosed for controlling initiation of a defrost cycle of an evaporator heat exchanger of a refrigeration system operatively associated with a refrigerated transport cargo box. The method includes the steps of establishing an return air-saturation temperature differential equal to the difference of a sensed air temperature of an air flow returning from the cargo box to pass over the heat exchange surface of the evaporator heat exchanger minus a refrigerant saturation temperature of a flow of refrigerant passing through the evaporator heat exchanger, comparing the return air-saturation temperature differential to a set point threshold defrost temperature differential, and if the return air-saturation temperature differential exceeds the set point threshold defrost temperature differential, initiating a defrost cycle for defrosting the evaporator heat exchanger.

CROSS REFERENCE TO RELATED APPLICATION

This application claims priority to U.S. Provisional Patent ApplicationSer. No. 61/360,651 entitled “Evaporator Refrigerant Saturation DemandDefrost,” filed on Jul. 1, 2010. The content of this application isincorporated herein by reference in its entirety.

FIELD OF THE INVENTION

This invention relates generally to refrigeration systems and, moreparticularly, to refrigerant vapor compression system evaporator coildefrost control and, more specifically, to the on demand initiation of adefrost cycle for the evaporator coil in response to a differentialbetween return air temperature and evaporator refrigerant saturationtemperature.

BACKGROUND OF THE INVENTION

Refrigerant vapor compression systems are well known in the art andcommonly used for conditioning air to be supplied to a climatecontrolled comfort zone within a residence, office building, hospital,school, restaurant or other facility. Refrigerant vapor compressionsystems are also commonly used in refrigerating air supplied to displaycases, merchandisers, freezer cabinets, cold rooms or otherperishable/frozen product storage area in commercial establishments.Refrigerant vapor compression systems are also commonly used intransport refrigeration systems for refrigerating air supplied to atemperature controlled cargo space of a truck, trailer, container or thelike for transporting perishable/frozen items by truck, rail, ship orintermodal.

Refrigerant vapor compression systems typically include a compressor, acondenser, an evaporator, and an expansion device. These basiccomponents are interconnected by refrigerant lines in a closedrefrigerant circuit, arranged in accord with known refrigerant vaporcompression cycles. The expansion device is disposed upstream, withrespect to refrigerant flow, of the evaporator and downstream of thecondenser. The evaporator includes a heat exchanger, typically a heatexchange tube coil, finned or un-finned, through which refrigerantflowing through the refrigerant circuit passes in heat exchangerelationship with air drawn from and circulated back to a temperaturecontrolled space. Because the air within the temperature controlledspace will contain moisture, to varying degrees, whether the climatecontrolled be an air conditioned room, a refrigerated display case, or atemperature controlled transport cargo box, and because the temperatureof the refrigerant flowing through the evaporator heat exchange tubecoil may drop below the freezing point of water, in some applicationsand under certain operating conditions, moisture in the air flowing overthe heat exchange tube coil will condense on the heat exchange surfaceof the tube coil and form frost. As the frost accumulates over time ofsystem operation in a cooling mode, the frost builds up on heat exchangesurface of the tube coil, adversely impacting heat transfer performanceand restricting air flow over the tube coil.

Consequently, it is customary practice to periodically interrupt systemoperation in a cooling mode and enter a defrost mode wherein theaccumulated frost is melted off the evaporator tube coil. A defrostcycle can be accomplished by reversing the flow of refrigerant throughthe refrigerant circuit so as to circulate a heated refrigerant,typically hot refrigerant vapor, through the evaporator heat exchanger.Defrost may also be accomplished through the activation of one or moreelectrical resistance heater operatively associated with the evaporatorheat exchange tube coil for heating the tube coil.

In operating refrigerant vapor compression systems, knowing when tointerrupt a cooling cycle to initiate a defrost cycle is important tooperating the refrigerant vapor compression system in a most efficientmanner. Initiating a defrost cycle at the expiration of specified timeintervals of operation in the cooling mode is a simple, but inefficient,control method. U.S. Pat. No. 6,205,800 discloses a method fordefrosting on demand by initiating a defrost routine for removingcondensate from an evaporator of a refrigerated device if the differencebetween the sensed air temperature within the refrigerated enclosure ofthe refrigerated device and the refrigerant temperature sensed by arefrigerant temperature sensor mounted on or disposed within theevaporator tube coil is greater or equal to a defrost threshold. U.S.Pat. No. 6,318,095 discloses controlling an outdoor coil defrost cycleon a reversible heat pump by continuously monitoring the differencebetween the outdoor coil temperature and the outdoor air temperature andinitiating a defrost cycle when that difference exceeds a target value.

Refrigerant vapor compression systems used in connection with transportrefrigeration systems are generally subject to more stringent operatingconditions due to the wide range of refrigeration load conditions andthe wide range of outdoor ambient conditions over which the refrigerantvapor compression system must operate to maintain product within thecargo space at a desired temperature. The refrigerant vapor compressionsystem must not only have sufficient capacity to rapidly pull down thetemperature of product loaded into the cargo space at ambienttemperature, but also should operate energy efficiently over the entireload range, including at low load when maintaining a stable producttemperature during transport.

The air within the transport cargo box may have a particularly highmoisture level after product is first loaded, therefore frost formationcan be particularly troublesome during pull down when maximum coolingcapacity is needed to draw down the product temperature as quickly aspossible. Excessive accumulation of frost on the evaporator tube coilresults in reduced heat transfer, which prolongs the time required forpull down. A common method currently in use in truck trailerapplications for controlling initiation of a defrost cycle relies on adifferential pressure switch that triggers a defrost cycle whenever theairside pressure drop across the evaporator tube coil exceeds a presetthreshold.

However, other factors that are not related to frost formation can alsoimpact the airside pressure drop. For example, field installed airchutes can significantly alter the air flow patterns and low airside airflows through the evaporator may not be sufficient to cause the pressuredifferential switch to trigger despite excessive frost formation of theheat exchange surface of the evaporator tube coil. Further, when thesystem is operating an low fan speeds, for example such as during astable temperature maintenance cooling mode or a low noise operationalmode, the airside air flow through the evaporator may again be too lowto cause the differential pressure switch to rigger despite excessivefrost accumulation on the evaporator tube coil.

Additionally, non-uniform frost/ice buildup is not an uncommon problemwith respect to evaporators associated with refrigerant vaporcompression systems in transport refrigeration applications. As a resultof air flow maldistribution through the evaporator heat exchanger,frost/ice buildup may be heavy on some sections of the evaporator heatexchange surface and nearly non-existent on other sections of theevaporator heat exchange surface. The air flow over the heat transfersurface becomes restricted and may not generate enough pressure drop totrigger an air pressure defrost switch to defrost the sections of theevaporator that have heavy frost/ice accumulations. Typically, intransport refrigeration applications, the refrigeration unit is providedwith a safety defrost which is automatically triggered whenever thetemperature differential between the sensed return air temperature and asensed evaporator heat exchanger surface temperature exceeds apreselected threshold, which is indicative of insufficient heated beingabsorbed by the refrigerant due to frost build-up on the evaporator heatexchange surface. The sensed surface temperature is typically taken by athermister mounted to the heat exchanger tube sheet or a tube fin, butcould also be mounted on the surface of a tube.

Continued cooling operation with an excessively frosted coil isinefficient. Cooling capacity may roll-off by 75% or more over as littleas two or three hours of operation in the cooling mode with anexcessively frosted coil. Continued cooling operation with anexcessively frosted coil also results in increased diesel fuelconsumption to power the refrigeration unit. Therefore, an active andmore direct method for initiating a defrost cycle that is directlyinfluenced by the build-up of frost on the heat exchange surface of theevaporator tube coil is needed.

SUMMARY OF THE INVENTION

A method is provided for controlling initiation of a defrost cycle of anevaporator heat exchanger of a refrigeration system. The method includesthe steps of: establishing a return air-saturation temperaturedifferential equal to the difference of a sensed air temperature of anair flow returning from the cargo box to pass over the evaporator heatexchanger minus a refrigerant saturation temperature of a flow ofrefrigerant passing through the evaporator heat exchanger; comparing thereturn air-saturation temperature differential to a set point thresholddefrost temperature differential; and if the return air-saturationtemperature differential exceeds the set point threshold defrosttemperature differential, initiating a defrost cycle for defrosting theevaporator heat exchanger.

The method may include the further step of sensing the air temperatureof and generating a signal indicative of the sensed air temperature ofan air flow returning from the cargo box to pass over the evaporatorheat exchanger. In an aspect, the method may further include the stepsof: sensing a refrigerant pressure of a flow of refrigerant passingthrough the evaporator heat exchanger and generating a signal indicativeof the sensed refrigerant pressure of the flow or refrigerant passingthrough the evaporator heat exchanger; and determining the refrigerantsaturation temperature based upon the sensed refrigerant pressuresignal.

In an aspect, the method may include the further steps of: sensing arefrigerant pressure of and generating a signal indicative of the sensedrefrigerant pressure of a flow of refrigerant passing through theevaporator heat exchanger at a plurality of spaced time intervals over aselected time period; calculating a plurality of refrigerant saturationtemperatures, one per each one of the plurality of refrigerant pressuressensed over the selected time period; calculating an adjustedrefrigerant saturation temperature based on the plurality of refrigerantsaturation temperatures; and establishing the return air-saturationtemperature differential as the difference of the sensed air temperatureminus the adjusted refrigerant saturation temperature. The step ofcalculating an adjusted refrigerant saturation temperature based on theplurality of refrigerant saturation temperatures may include calculatingthe adjusted refrigerant saturation temperature as an arithmetic averageof the plurality of refrigerant saturation temperatures. The step ofcalculating an adjusted refrigerant saturation temperature based on theplurality of refrigerant saturation temperatures may include calculatingthe adjusted refrigerant saturation temperature as an arithmetic mean ofthe plurality of refrigerant saturation temperatures. In an aspect, theselected time period may range from at least about three minutes up toabout five minutes.

In an aspect, the method may include the further step of adjusting theset point threshold defrost temperature differential as a function ofrefrigerant mass flow rate of the refrigerant flowing through theevaporator heat exchanger prior to comparing the return air-saturationtemperature differential to the set point threshold defrost temperaturedifferential. In an aspect, the method may include the further steps of:calculating a clean coil temperature differential equal to thedifference of the sensed return air temperature minus the refrigerantsaturation temperature following termination of the defrost cycle;resetting the set point threshold defrost temperature to be the cleancoil temperature differential plus a predetermined temperature delta;and initiating the next defrost cycle when the return air-saturationtemperature differential exceeds the reset set point temperaturedifferential.

BRIEF DESCRIPTION OF THE DRAWINGS

For a further understanding of the disclosure, reference will be made tothe following detailed description which is to be read in connectionwith the accompanying drawing, wherein:

FIG. 1 is a perspective view of a truck trailer equipped with arefrigeration unit operatively associated with a temperature controlledcargo box;

FIG. 2 is a schematic diagram of an exemplary embodiment of arefrigerant vapor compression system associated with the refrigerationunit of the truck trailer of FIG. 1;

FIG. 3 is a schematic illustration of an exemplary embodiment of theevaporator heat exchanger of the refrigerant vapor compression system ofFIG. 2; and

FIG. 4 is a block diagram illustrating an exemplary embodiment of themethod disclosed herein.

FIG. 5 is a block diagram illustrating an alternate embodiment of themethod illustrated in FIG. 4; and

FIG. 6 is a block diagram illustrating an additional step to the methodillustrated in FIG. 4.

DETAILED DESCRIPTION OF THE INVENTION

Referring initially to FIG. 1, there is shown a truck trailer 100 havinga refrigerated cargo box 110 having access doors 112 that open to theinterior space 114 of the cargo box from the exterior of the trucktrailer to facilitate loading of product into the cargo box 110 fortransport. The truck trailer 100 is equipped with a transportrefrigeration unit 10 for regulating and maintaining a temperaturecontrolled atmosphere within the cargo box during transport within adesired storage temperature range selected for the perishable productbeing shipped therein. Although the demand defrost method disclosedherein will be described herein with reference to the refrigerated cargobox of the depicted truck trailer, it is to be understood that theinvention may also be used in connection with other refrigerated cargotransport boxes, including for example a refrigerated box of a truck, ora refrigerated cargo container for transporting perishable product byship, by rail, by road or intermodal transport. The disclosed demanddefrost method may also be applied to controlling evaporator defrostcycle initiation on demand in refrigerant vapor compression systems forsupplying conditioned air to a temperature controlled space, such as usein connection with air conditioning systems and commercial refrigerationsystems.

Referring now also to FIG. 2, the transport refrigeration unit 10includes a refrigerant vapor compression system 12 and an associatedpower source. The refrigerant vapor compression system 12 includes acompression device 20, a condenser 30 having a heat exchanger andassociated condenser fan(s) 34, an evaporator 40 having a heat exchanger42 and associated evaporator fan(s) 44, and an evaporator expansiondevice 46, all arranged in a conventional refrigeration cycle andconnected in a refrigerant circulation circuit including refrigerantlines 22, 24, 26 and the condenser tubular heat exchanger 32 and theevaporator tubular heat exchanger 42. The transport refrigeration system10 is mounted as in conventional practice to an exterior wall of thetruck trailer 100, for example the front wall 116 thereof, with thecompressor 20 and the condenser 30 with its associated condenser fan(s)34 and power source 50 disposed externally of the refrigerated cargo box110 in a housing 118.

The evaporator 40 extends through an opening in the front wall 116 intothe refrigerated cargo box 110. The expansion device 46, which in thedepicted embodiment is an electronic expansion valve, but could be athermostatic expansion valve, is disposed in refrigerant line 24downstream with respect to refrigerant flow of the condenser heatexchanger 32 and upstream with respect to refrigerant flow of theevaporator heat exchanger 42 for metering the flow of refrigerantthrough the evaporator in response to the degree of superheat in therefrigerant at the outlet of the evaporator 40, as in conventionalpractice. A refrigerant pressure sensor 48 is mounted on the tubularheat exchanger 42 of the evaporator 40 for monitoring the sensing therefrigerant flowing through the evaporator heat exchanger 42 at or nearthe outlet thereof. Although the particular type of evaporator heatexchanger 42 used is not limiting of the invention, the evaporator heatexchanger 42 may, for example, comprise one or more heat exchange tubecoils, as depicted in the drawing, or one or more tube banks formed of aplurality of tubes extending between respective inlet and outletmanifolds. The tubes may be round tubes or flat tubes and may be finnedor un-finned.

The compressor 20 may comprise a single-stage or multiple-stagecompressor such as, for example, a reciprocating compressor or a scrollcompressor, although the particular type of compressor used is notgermane to or limiting of the invention. In the exemplary embodimentdepicted in FIG. 2, the compressor is a reciprocating compressor, suchas for example, an 06D model reciprocating compressor manufactured byCarrier Corporation or a variant thereof, having a compressingmechanism, an internal electric compressor motor and an interconnectingdrive shaft that are all sealed within a common housing of thecompressor 20. The power source 50 powers the internal electric motor ofthe compressor. In an embodiment, the power source 50 generatessufficient electrical power for fully driving the electrical motor ofthe compressor 20 and also for providing all other electrical powerrequired by the fans 34, 44 and other parts of the refrigeration unit10. In an electrically powered embodiment of the transport refrigerationunit 10, the power source 50 comprises a single on-board engine drivensynchronous generator configured to selectively produce at least one ACvoltage at one or more frequencies. An electrically powered transportrefrigeration system suitable for use on truck trailer transportvehicles are shown in U.S. Pat. No. 6,223,546, assigned to the assigneeof the present application, the entire disclosure of which isincorporated herein by reference.

The transport refrigeration unit 10 also includes an electroniccontroller 60 that is configured to operate the transport refrigerationunit 10 to maintain a predetermined thermal environment within theinterior space 114 defined within the cargo box 110 wherein the productis stored during transport. The electronic controller 60 maintains thepredetermined thermal environment by selectively powering andcontrolling the operation of various components of the refrigerant vaporcompression system, including the compressor 20, the condenser fan(s) 34associated with the condenser 30, the evaporator fan(s) 44 associatedwith the evaporator 40, and various valves in the refrigerant circuit,including but not limited to the electronic expansion valve 46 (ifpresent) and the suction modulation valve 62 (if present). When coolingof the environment within interior space 114 of the cargo box 110 isrequired, the electronic controller 60 activates the compressor 20, thecondenser fan(s) 34 and the evaporator fan(s) 44, as appropriate, andadjusts the position of the electronic expansion valve 46 to meter theflow of refrigerant through the evaporator heat exchanger 42 to providea desired degree of superheat in the refrigerant vapor at the evaporatoroutlet, and adjusts the position of the suction modulation valve 62 toincrease or decrease the flow of refrigerant supplied to the compressor20 as appropriate to control and stabilize the temperatures within theinterior space 114 within the cargo box 110 at the respective set pointthreshold defrost temperature, which corresponds to the desired productstorage temperatures for the particular products stored within cargo box110.

In one embodiment, the electronic controller 60 includes amicroprocessor and an associated memory. The memory of the controller 60may be programmed to contain preselected operator or owner desiredvalues for various operating parameters within the system, including,but not limited to, a temperature set point for the air within theinterior space 114 of the cargo box 110, refrigerant pressure limits,current limits, engine speed limits, and any variety of other desiredoperating parameters or limits within the system. The programming of thecontroller is within the ordinary skill in the art. The controller 60may include a microprocessor board that includes the microprocessor, anassociated memory, and an input/output board that contains ananalog-to-digital converter which receives temperature inputs andpressure inputs from a plurality of sensors located at various pointsthroughout the refrigerant circuit and the refrigerated cargo box,current inputs, voltage inputs, and humidity levels. The input/outputboard may also include drive circuits or field effect transistors andrelays which receive signals or current from the controller 60 and inturn control various external or peripheral devices associated with thetransport refrigeration system. In an embodiment, the controller 60 maycomprise the MicroLink™ controller available from Carrier Corporation,the assignee of this application. However, the particular type anddesign of the controller 60 is within the discretion of one of ordinaryskill in the art to select and is not limiting of the invention.

As in conventional practice, when the refrigerant vapor compressionsystem is in operation, low temperature, low pressure refrigerant vaporis compressed by the compressor 20 to a high pressure, high temperaturerefrigerant vapor and passed from the discharge outlet of the compressor20 into refrigerant line 22. The refrigerant circulates through therefrigerant circuit via refrigerant line 22 to and through the heatexchange tube coil or tube bank of the condenser heat exchanger 32,wherein the refrigerant vapor condenses to a liquid, and the subcooler32, thence through refrigerant line 24 through a first refrigerant passof the refrigerant-to-refrigerant heat exchanger 35, and thencetraversing the evaporator expansion device 46 before passing through theevaporator heat exchanger 42 and thence through refrigerant line 26,passing a second refrigerant pass of the refrigerant-to-refrigerant heatexchanger 35 before passing to the suction inlet of the compressiondevice 20.

In flowing through the heat exchange tube coil or tube bank of theevaporator heat exchanger 42, the refrigerant evaporates, and istypically superheated, as it passes in heat exchange relationship theair passing through the airside of the evaporator 40. The air is drawnfrom within the cargo box 110 by the evaporator fan(s) 44, passed overthe external heat transfer surface of the heat exchange tube coil ortube bank of the evaporator heat exchanger 42 and circulated back intothe interior space 114 of the cargo box 110. The air drawn from thecargo box 110 is referred to as “return air” and the air circulated backto the cargo box 110 is referred to as “supply air”. It is to beunderstood that the term “air’ as used herein includes mixtures of airand other gases, such as for example, but not limited to nitrogen orcarbon dioxide, sometimes introduced into a refrigerated cargo transportbox. A temperature sensor 45 is provided to sense the actual temperatureof the return air drawn from the temperature controlled interior space114 of the cargo box 110 before passing over the evaporator heatexchanger 42.

During operation of the refrigerant vapor compression system in acooling mode, moisture in the return air will condense onto the heattransfer surface, i.e. surface of the tubes and the fins if finned tubesare present, of the evaporator heat exchanger 42 as the return air iscooled in passing in heat exchange relationship with the refrigerantflowing through the evaporator heat exchanger 42. The condensate willfreeze on the heat transfer surface of the evaporator heat exchanger 42and tend to accumulate as a layer of frost and/or ice on the heattransfer surface of the evaporator heat exchanger 42. As the frost/icelayer builds-up, heat transfer performance of the evaporator heatexchanger 42 deteriorates and the airside flow area through theevaporator heat exchanger 42 becomes more and more restricted.Therefore, operation of the refrigerant vapor compression system in thecooling mode must be interrupted to conduct an evaporator defrost cyclewhenever the accumulated frost/ice layer becomes excessive.

Referring now to FIG. 3, an electrical resistance heater 70 is providedin operative association with the evaporator heat exchanger 42 to meltthe accumulated frost/ice deposited on the heat transfer surface of theevaporator heat exchanger 42. Whenever a defrost cycle is to beimplemented, the controller 60 will deactivate the compression device20, the condenser fan(s) 34 and the evaporator fan(s) 44 for theduration of the defrost cycle and activate the electrical resistanceheater 70 for the duration of the defrost cycle by selectively switchingon the supply of electrical power from power source 50.

The controller 60 will terminate the defrost cycle by deactivating, i.e.switching off the supply of electrical power to, the electricalresistance heater 70. The controller 60 may terminate the defrost cycleafter a predetermined period of time in operation in the defrost cycleelapses or may terminate the defrost cycle based on a temperature signalfrom a coil defrost termination sensor indicative of a sensed surfacetemperature indicative of an external tube surface temperature of theevaporator heat exchanger 42. After termination of the defrost cycle,the controller 60 will return the refrigerant vapor compression systemto operation in the cooling mode, by restarting the compression device20, the condenser fan(s) 34 and the evaporator fan(s) 44. Thus, duringdefrost cycle operation, not only is the air to the controlled space notbeing cooled, but the heat transfer surface of the evaporator heatexchanger 42 is also being heated.

Referring now to FIG. 4, in accord with the method disclosed herein, thecontroller 60 will initiate a defrost cycle based on an returnair-saturation temperature differential (RASTD), which is defined as theactual return air temperature (RAT), which is sensed by the return airtemperature sensor 45 at step 202, minus the refrigerant saturationtemperature (ERST) within the evaporator heat exchanger 42. Thecontroller 60 uses the signal indicative of the sensed return airtemperature generated by and received from the return air temperaturesensor 45 in controlling operation of the refrigeration unit in thecooling mode and also uses the signal indicative of the sensedevaporator refrigerant pressure (ERP) generated by and received from thepressure sensor 48 at step 204 to calculate the evaporator refrigerantsaturation temperature (ERST) for controlling the electronic expansionvalve 46 to control superheat. Additionally, in accord with an aspect ofthe method disclosed herein, the controller 60 will, at step 206,determine the evaporator refrigerant saturation temperature (ERST) basedupon the sensed evaporator refrigerant pressure (ERP) sensed by pressuresensor 48 at step 204, and at step 208 calculate the returnair-saturation temperature differential (RASTD) by subtracting theevaporator refrigerant saturation temperature (ERST) from the actualreturn air temperature (RAT) sensed by the return air temperature sensor45 at step 202.

The controller 60, at step 210, compares the calculated returnair-saturation temperature differential (RASTD) to a defrost thresholddefrost temperature differential (DTSP). If the calculated returnair-saturation temperature differential does not exceed the defrostthreshold approach temperature differential at block 212, the controller60 continues operation of the refrigerant vapor compression system inthe refrigeration (cooling) mode and repeats steps 202 through 210.However, if the calculated return air-saturation temperaturedifferential exceeds the defrost threshold defrost temperaturedifferential at block 214, the controller 60, interrupts operation ofthe refrigerant vapor compression system in the refrigeration (cooling)mode and initiates a defrost cycle to remove frost/ice accumulated onthe heat transfer surface of the evaporator heat exchanger 42 in themanner discussed hereinbefore. The controller 60 continues operation ofthe refrigerant vapor compression system 10 in the defrost cycle untilall or at least substantially all of the frost/ice accumulated on theheat transfer surface of the evaporator heat exchanger 42 has beenremoved.

Referring now to FIG. 5, in an aspect of the method depicted therein, atstep 207, the controller 60 may calculate an adjusted evaporatorrefrigerant saturation temperature as a function of a plurality ofinstantaneous evaporator refrigeration saturation temperatures (ERSi)sensed at spaced time intervals (step 206) to filter out evaporatorsuperheat control related noise and any influence on control logic. Inan embodiment, the controller 60 calculates the adjusted evaporatorrefrigerant saturation temperature as a running average of a pluralityof instantaneous evaporator refrigerant saturation temperatures over aselected period of time. In an embodiment, the controller 60 maycalculate the adjusted evaporator refrigerant saturation temperature asan arithmetic mean of a plurality of instantaneous evaporatorrefrigerant saturation temperatures over a selected period of time. Forexample, the adjusted evaporator refrigerant saturation temperature maybe calculated as the arithmetic average or arithmetic mean of thoseinstantaneous evaporator refrigerant saturation temperatures calculatedover the immediately past three to five minutes.

In an aspect of the method disclosed herein, the controller 60 maycompensate for variation in refrigerant mass flow rate through theevaporator heat exchanger 42 by adjusting the threshold defrosttemperature differential (TDTD) as a function of the refrigerant massflow rate. For example, the controller 60 may select the thresholddefrost return air-saturation temperature differential from aninitiation curve of threshold defrost return air-saturation temperaturedifferential versus refrigerant mass flow rate through the evaporatorheat exchanger 42. The initiation curve may be empirically developedbased on testing of the actual refrigerant vapor compression system inuse. In determining whether or not to initiate a defrost cycle, thecontroller 60 will compare the calculated return air-saturationtemperature differential to a adjusted threshold defrost temperaturedifferential selected from the aforementioned initiation curve based onthe actual refrigerant mass flow rate through the evaporator heatexchanger 42 associated with the evaporator refrigerant saturationtemperature used in calculating the return air-saturation temperaturedifferential. If the calculated return air-saturation temperaturedifferential comprises an adjusted return air-saturation temperaturedifferential based on a plurality of instantaneous return air-saturationtemperature differentials, then the evaporator refrigerant mass flowrate associated therewith for purposes of selection of the adjustedthreshold defrost temperature differential would be the correspondingaverage or mean evaporator refrigerant mass flow rate.

In a further aspect of the method disclosed herein, the thresholddefrost temperature differential may be selected based on a sensed“clean coil” return air-saturation temperature differential, Forexample, in implementing this aspect of the method, at the terminationof each defrost cycle when the heat exchange surface of the evaporatorheat exchanger 42 is substantially frost/ice free, the controller 60will calculate a “clean coil” return air-saturation temperaturedifferential based upon the then current sensed return air temperatureand evaporator refrigerant saturation temperature. The controller 60would then set the defrost threshold approach temperature differentialfor triggering the next defrost cycle to be a pre-determined temperaturedelta from that “clean coil” return air-saturation temperaturedifferential. Thus, to trigger a defrost cycle, the returnair-saturation temperature differential would need to exceed the actual“clean coil” return air-saturation temperature differential attermination of the last previous defrost cycle by a pre-determinedtemperature delta. In this aspect of the method disclosed herein, theinitiation of defrost cycles is automatically adapted in response tooperating conditions associated with the particular product beingshipped, local ambient conditions, loading, air flow variations, andother operational factors that may potentially influence frost/iceformation.

The method for initiating a defrost cycle as discloses relies oninformation available from conventional sensors that are customarilyprovided on conventional refrigerant vapor compression systems andtherefore does not require the installation of new hardware.Additionally, the method disclosed herein eliminates the need for an airpressure switch for initiating defrost, thereby reducing cost andimproving overall reliability. Further, triggering defrost based onreturn air-saturation temperature differential in accord with the methoddisclosed herein, allows for more effective and more efficient coolingoperation by reducing unnecessary run time in the cooling with a highlyfrosted evaporator while waiting for a safety type defrost to initiatebecause the air pressure switch failed to trigger a defrost cycle whenneeded.

As frost builds up on the tube coil or tube bank of the evaporator heatexchanger 42, the air flow through the evaporator 40 goes down theairside pressure drop increases. Consequently, the refrigerant flowingthrough the heat exchanger tubes absorbs less heat. Therefore, withoutas much heat going into the refrigerant, the expansion valve 46throttles the refrigerant flow passing through the tubes of theevaporator heat exchanger 42 in an attempt to maintain the desiredrefrigerant superheat, which results in a drop in evaporator refrigerantpressure. Thus, the refrigerant saturation temperature also decreases.As the refrigerant saturation temperature goes lower and lower a s theexpansion valve continues to throttle the refrigerant flow, thetemperature differential with respect to the sensed return airtemperature increases, which will lead to a demand defrost when thethreshold defrost temperature differential is exceeded. However, a lowrefrigerant pressure condition resulting from a low refrigerant flow inthe evaporator despite a wide open (for example 90% or more open), whichcould result from a loss of refrigerant charge, could result in an ondemand defrost cycle being called for when the frost build-up per sedoes not warrant a defrost. Referring now to FIG. 6, to avoid theinitiation of an on demand defrost cycle, the controller 60, at step216, will monitor the position of the expansion valve 46 and the degreeof superheat in the refrigerant leaving the evaporator heat exchanger 42as a feedback to detect whether the one demand defrost indication is theresult of a low refrigerant flow condition through the evaporator heatexchanger 42 rather than the result of excessive frost build-up. Thecontroller 60 will determine whether both the position of the expansionvalve 46 and the superheat are within normal operating range. If so, thecontroller 60 will terminate operation in the cooling mode and initiatea defrost cycle. If not, the controller will continue operation in thecooling mode.

The terminology used herein is for the purpose of description, notlimitation. Specific structural and functional details disclosed hereinare not to be interpreted as limiting, but merely as basis for teachingone skilled in the art to employ the present invention. Those skilled inthe art will also recognize the equivalents that may be substituted forelements described with reference to the exemplary embodiments disclosedherein without departing from the scope of the present invention.

While the present invention has been particularly shown and describedwith reference to the exemplary embodiment as illustrated in thedrawing, it will be recognized by those skilled in the art that variousmodifications may be made without departing from the spirit and scope ofthe invention. Therefore, it is intended that the present disclosure notbe limited to the particular embodiment(s) disclosed as, but that thedisclosure will include all embodiments falling within the scope of theappended claims.

We claim:
 1. A method for controlling initiation of a defrost cycle ofan evaporator heat exchanger of a refrigerant vapor compression systemfor supplying conditioned air to a temperature controlled space, themethod comprising the steps of: establishing a return air-saturationtemperature differential equal to the difference of a sensed airtemperature of an air flow returning from the temperature controlledspace to pass over the evaporator heat exchanger minus a refrigerantsaturation temperature of a flow of refrigerant passing through theevaporator heat exchanger; comparing the return air-saturationtemperature differential to a set point threshold defrost temperaturedifferential; and if the return air-saturation temperature differentialexceeds the set point threshold defrost temperature differential,initiating a defrost cycle for defrosting the evaporator heat exchanger.2. The method as recited in claim 1 further comprising the step ofsensing the air temperature of and generating a signal indicative of thesensed air temperature of an air flow returning from the temperaturecontrolled space to pass over the evaporator heat exchanger.
 3. Themethod as recited in claim 1 further comprising the steps of: sensing arefrigerant pressure of and generating a signal indicative of the sensedrefrigerant pressure of a flow of refrigerant passing through theevaporator heat exchanger; determining the refrigerant saturationtemperature based upon the sensed refrigerant pressure signal.
 4. Themethod as recited in claim 1 further comprising the steps of: sensing arefrigerant pressure of and generating a signal indicative of the sensedrefrigerant pressure of a flow of refrigerant passing through theevaporator heat exchanger at a plurality of spaced time intervals over aselected time period; calculating a plurality of refrigerant saturationtemperatures, one per each one of the plurality of refrigerant pressuressensed over the selected time period; calculating an adjustedrefrigerant saturation temperature based on the plurality of refrigerantsaturation temperatures; and establishing the return air-saturationtemperature differential as the difference of the sensed air temperatureminus the adjusted refrigerant saturation temperature.
 5. The method asrecited in claim 4 wherein the step of calculating an adjustedrefrigerant saturation temperature based on the plurality of refrigerantsaturation temperatures comprises calculating the adjusted refrigerantsaturation temperature as an arithmetic average of the plurality ofrefrigerant saturation temperatures.
 6. The method as recited in claim 4wherein the step of calculating an adjusted refrigerant saturationtemperature based on the plurality of refrigerant saturationtemperatures comprises calculating the adjusted refrigerant saturationtemperature as an arithmetic mean of the plurality of refrigerantsaturation temperatures.
 7. The method as recited in claim 4 wherein theselected time period ranges from at least about three minutes up toabout five minutes.
 8. The method as recited in claim 1 furthercomprising the step of adjusting the set point threshold defrosttemperature differential as a function of refrigerant mass flow rate ofthe refrigerant flowing through the evaporator heat exchanger prior tocomparing the return air-saturation temperature differential to the setpoint threshold defrost temperature differential.
 9. The method asrecited in claim 1 further comprising the steps of: calculating a cleancoil temperature differential equal to the difference of the sensedreturn air temperature minus the refrigerant saturated temperaturefollowing termination of the defrost cycle; resetting the set pointthreshold defrost temperature differential to be the clean coiltemperature differential plus a predetermined temperature delta; andinitiating the next defrost cycle when the return air-saturationtemperature differential exceeds the reset set point threshold defrosttemperature differential.
 10. The method as recited in claim 1 furthercomprising the step of determining that the position of an evaporatorexpansion valve is within normal operating range prior to initiating ademand defrost.
 11. A method for controlling initiation of a defrostcycle of an evaporator heat exchanger of a refrigeration systemoperatively associated with a refrigerated transport cargo box, themethod comprising the steps of: establishing a return air-saturationtemperature differential equal to the difference of a sensed airtemperature of an air flow returning from the cargo box to pass over theevaporator heat exchanger minus a refrigerant saturation temperature ofa flow of refrigerant passing through the evaporator heat exchanger;comparing the return air-saturation temperature differential to a setpoint threshold defrost temperature differential; and if the returnair-saturation temperature differential exceeds the set point thresholddefrost temperature differential, initiating a defrost cycle fordefrosting the evaporator heat exchanger.
 12. The method as recited inclaim 11 further comprising the step of sensing the air temperature ofand generating a signal indicative of the sensed air temperature of anair flow returning from the cargo box to pass over the evaporator heatexchanger.
 13. The method as recited in claim 11 further comprising thesteps of: sensing a refrigerant pressure of and generating a signalindicative of the sensed refrigerant pressure of a flow of refrigerantpassing through the evaporator heat exchanger; determining therefrigerant saturation temperature based upon the sensed refrigerantpressure signal.
 14. The method as recited in claim 11 furthercomprising the steps of: sensing a refrigerant pressure of andgenerating a signal indicative of the sensed refrigerant pressure of aflow of refrigerant passing through the evaporator heat exchanger at aplurality of spaced time intervals over a selected time period;calculating a plurality of refrigerant saturation temperatures, one pereach one of the plurality of refrigerant pressures sensed over theselected time period. calculating an adjusted refrigerant saturationtemperature based on the plurality of refrigerant saturationtemperatures; and establishing the return air-saturation temperaturedifferential as the difference of the sensed air temperature minus theadjusted refrigerant saturation temperature.
 15. The method as recitedin claim 14 wherein the step of calculating an adjusted refrigerantsaturation temperature based on the plurality of refrigerant saturationtemperatures comprises calculating the adjusted refrigerant saturationtemperature as an arithmetic average of the plurality of refrigerantsaturation temperatures.
 16. The method as recited in claim 14 whereinthe step of calculating an adjusted refrigerant saturation temperaturebased on the plurality of refrigerant saturation temperatures comprisescalculating the adjusted refrigerant saturation temperature as anarithmetic mean of the plurality of refrigerant saturation temperatures.17. The method as recited in claim 14 wherein the selected time periodranges from at least about three minutes up to about five minutes. 18.The method as recited in claim 11 further comprising the step ofadjusting the set point threshold defrost temperature differential as afunction of refrigerant mass flow rate of the refrigerant flowingthrough the evaporator heat exchanger prior to comparing the returnair-saturation temperature differential to the set point temperaturedifferential.
 19. The method as recited in claim 11 further comprisingthe steps of: calculating a clean coil temperature differential equal tothe difference of the sensed return air temperature minus therefrigerant saturated temperature following termination of the defrostcycle; resetting the set point threshold defrost temperature to be theclean coil temperature differential plus a predetermined temperaturedelta; and initiating the next defrost cycle when the returnair-saturation temperature differential exceeds the reset set pointthreshold defrost temperature differential.
 20. The method as recited inclaim 11 further comprising the step of determining that the position ofan evaporator expansion valve is within normal operating range prior toinitiating a demand defrost.